Multiple-clutch device

ABSTRACT

The invention is directed to a multiple-clutch device, such as a double-clutch device, for arranging in a drivetrain of a motor vehicle between a drive unit and a transmission, wherein the clutch device has a first clutch arrangement associated with a first transmission input shaft of the transmission and a second clutch arrangement associated with a second transmission input shaft of the transmission for transmitting torque between the drive unit and the transmission. According to one aspect of the invention, it is suggested that at least one of the clutch arrangements has an actuating piston defining a pressure chamber for actuation, preferably for engagement of the clutch arrangement by means of a pressure medium, preferably a hydraulic medium, wherein the actuating piston divides the pressure chamber from an associated centrifugal force pressure compensation chamber which receives a pressure compensation medium.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention is directed to double-clutch device, for arranging in adrivetrain of a motor vehicle between a drive unit and a transmission,wherein the clutch device has a first clutch arrangement associated witha first transmission input shaft of the transmission and a second clutcharrangement associated with a second transmission input shaft of thetransmission for transmitting torque between the drive unit and thetransmission.

2. Description of the Related Art

A clutch device of this type is known, for example, from EP 0 931 951A1. The clutch device serves to connect the drive of a motor vehiclewith a multiple-speed shift transmission via two friction clutches whichare preferably automatically actuated, wherein a disengagement orrelease system is allocated to each of these two friction clutches, sothat the two friction clutches can be engaged or released independentlyfrom one another. A clutch disk of one of the two friction clutches isarranged on a central transmission input shaft so as to be fixed withrespect to rotation relative to it, while a clutch disk of the otherfriction clutch engages at a second transmission input shaft so as to befixed with respect to rotation relative to it, this second transmissioninput shaft, constructed as a hollow shaft, enclosing the centraltransmission input shaft. The known double-clutch is arranged with afixed thrust plate of one friction clutch at a flywheel of an internalcombustion engine. To this extent, the arrangement of the double-clutchin a drivetrain substantially corresponds to the arrangement ofconventional (single-)friction clutches in the drivetrain.

Double-clutch devices (called simply double-clutches) of the typementioned above have attracted great interest recently and are generallyformed of two wet or dry clutches which are switched alternately,possibly also with overlapping. Particularly in connection with amultiple-speed shift transmission, clutches of this type make itpossible to carry out shifting processes between two respectivetransmission speeds of the transmission without interruption of tractiveforces.

In principle, double-clutch devices make it possible for both clutchesto be applied jointly in especially difficult starting processes,particularly those common in car racing. For this purpose, theaccelerator pedal can be deflected to its stop, as the case may be,while the motor vehicle is kept essentially stationary at the same timeby applying the maximum braking force until the clutch has reached itsoptimal transmission point. When the braking action is canceled at themoment of reaching the optimal transmission point, the vehicle isstarted with maximum acceleration. Starting processes of this kind arealso considered for motor vehicles with a relatively weak engine underextreme starting conditions, for example, when starting on anobstruction; that is, they are not considered only for racing cars.

Obviously, starting processes of the type described above lead to highslippage with a correspondingly extensive development of heat. Thispresents the problem of carrying away this heat from the area of thefriction clutch serving as starting clutch. Further, a correspondinglyhigh wear of the friction clutch must be taken into account. Moreover,heating of the friction clutches is accompanied by changes in thecoefficient of friction of the friction clutches, so that control of therelease mechanisms of the two friction clutches, and therefore controlof the two friction clutches relative to one another, can be appreciablyimpaired. Since inaccuracies or changes in the functional matching ofthe two friction clutches relative to one another caused by heat canhave the result that a torque ratio not intended in the shifting processis applied to the transmission input shafts, shifting processes in theshift transmission can be subjected to load. The synchronization in theshift transmission can be overtaxed in this way, so that, in the worstcase, the shift transmission can be damaged to the point of completefailure, apart from disadvantages with respect to efficiency which occurin any case. On the whole, mismatching between the two friction clutchescaused by heat is incompatible with a problem-free torque transmissionin shifting processes in the shift transmission without interruption oftractive force and without jerking during shifting.

Another problem area in a double-clutch device relates to startingprocesses carried out in opposition to an inclination, wherein the motorvehicle must be prevented from rolling backward, or those which are usedwhen parking at the lowest possible speed, for example, for precisepositioning of a motor vehicle in a parking space. The operating statesmentioned above are referred to in technical circles as “hill-holding”and “creeping”. Both starting processes have in common that the frictionclutch serving as starting clutch is operated, sometimes withoutactuation of the accelerator, over a longer period of time with slip.Although the torques to be transmitted in such starting processes liewell below those occurring under the operating conditions describedabove, especially in car racing, an intensive heating of the respectivefriction clutch or even both friction clutches can occur, resulting inthe problems detailed above.

Suggestions have been made for gear-shifting strategies and shiftingprocesses for double-clutch transmissions based on the aimed foradjustment of clutch slip (DE 196 31 983 C1) with consequent generationof friction heat. Depending on driving behavior, overheating problems ofthe type mentioned above cannot be ruled out.

The risk of intensive overheating exists not only in a dry frictionclutch, but can also occur in so-called “wet” friction clutches, in theform of a disk or plate clutch, which are operated by the action of aviscous operating medium such as hydraulic fluid. By way of example, agear change box with two plate clutches is known from DE 198 00 490 A1,wherein one plate clutch is provided for forward driving and the otherfor driving in reverse. DE 198 00 490 A1 is concerned primarily withproviding adequate cooling of the two plate clutches using the viscousoperating medium. In spite of the liquid cooling, heating of thefriction clutches is also a considerable problem in plate clutchesbecause the operating medium, which usually flows through frictionfacing grooves or the like to carry off the heat, cannot be guidedbetween the plates in optional quantity. The reason for this is that, onone hand, excessive flow through the friction facing grooves or the likewould build up a counterpressure between the friction surfaces of twoadjacent plates and would therefore reduce the capacity of the frictionclutches to transmit torque (with a corresponding increase in slip andtherefore additional generation of friction heat, so that the problem ofoverheating is exacerbated) and, on the other hand, the operating mediumcould be overheated and destroyed when flowing through between theplates. Overheating in plate clutches can result in that the frictionsurfaces can no longer separate from one another completely during adisengaging process and, consequently, torques can still be transmittedvia the clutch which should be disengaged, so that considerable dragtorques can reach the associated shift transmission. When platesclutches are used in a multiple-clutch device, especially adouble-clutch device, of the type mentioned above, shifting processescould again be brought under load in the shift transmission withresulting overtaxing of the synchronization in the shift transmission.

One approach to mastering overheating problems in the area of frictionclutches in case of unfavorable operating conditions, for example, withproblematic starting processes in a motor vehicle, is to provide anotherstarting element in addition to the first and second clutch arrangementswhich is in the form of a hydraulic clutch or hydrodynamic clutch andcomprises a hydrodynamic circuit with an impeller wheel, a turbine wheeland, if desired, a stator wheel. The driving member can be connected inparallel with one of the two friction clutches; that is, it can act on acommon transmission input shaft irrespective of the engagement state ofthis friction clutch. A clutch device in which two plate clutches and astarting element of this type are integrated, was described in theGerman Patent Application 199 46 857.5 by the present Applicant whichwas applied for on Sep. 30, 1999 and whose disclosure is incorporated inthe subject matter disclosed in the present application.

Within the framework of investigations undertaken by the presentApplicants in connection with double-clutch devices, it was shown ingeneral that wet-type clutches exhibit sealing problems and problemsrelating to output losses. Further, it was shown that boundaryconditions relating to the available axial and radial installation spacecould be adhered to only with difficulty, if at all, based on thepreviously known concepts. With regard to clutches, possibly, diaphragmclutches, which are actuated by pistons integrated in the clutch device,the arrangement of the piston chambers associated with the pistonsproved especially problematic.

SUMMARY OF THE INVENTION

The object of the invention is generally to achieve improvements withrespect to at least one of the problems and/or other problems mentionedabove. In particular, an objective of the invention is to provide amultiple-clutch device with advantageous operating performance, forexample, with respect to operating reliability, and/or a multiple-clutchdevice which can be constructed in a compact manner.

According to a first aspect, the invention provides a multiple-clutchdevice, i.e., a double-clutch device, for arranging in a drivetrain of amotor vehicle between a drive unit and a transmission, wherein theclutch device has a first clutch arrangement associated with a firsttransmission input shaft of the transmission and a second clutcharrangement associated with a second transmission input shaft of thetransmission for transmitting torque between the drive unit and thetransmission. According to this first aspect of the invention, at leastone of the clutch arrangements has an actuating piston defining apressure chamber for actuation, preferably for engagement of the clutcharrangement by means of a pressure medium, preferably a hydraulicmedium, wherein the actuating piston divides the pressure chamber froman associated centrifugal force pressure compensation chamber holding apressure compensation medium which compensates for increases in pressurein the pressure chamber caused by centrifugal force and, consequently,for forces on the actuating piston caused by centrifugal force.

The centrifugal force pressure compensation chamber can be connected toa pressure compensation medium supply in order to supply pressurecompensation medium to the centrifugal force pressure compensationchamber at least in a requested operating state of the clutch device.The pressure compensation medium supply can be formed by a hydraulicmedium supply or a separate operating fluid supply, possibly anoperating oil supply, which also serves to supply at least one otherfunctional unit of the clutch unit.

When a return spring or restoring spring arrangement is provided forrestoring the actuating piston, this restoring spring arrangement can bereceived in the centrifugal force pressure compensation chamber in orderto economize on installation space.

It is particularly preferable that a sealing arrangement associated withthe pressure compensation chamber and/or a boundary wall of the pressurecompensation chamber are/is so arranged and so constructed in an elasticmanner that they/it at least reinforce/reinforces a disengagement of theclutch arrangement in a stroke area of the actuating piston comprising astroke position of the actuating piston corresponding to an engagedstate of the respective clutch arrangement. The sealing arrangementand/or boundary wall can accordingly assume the function of a springelement which reinforces or brings about the opening of the actuatingpiston. As a rule, the above-mentioned restoring spring arrangement isadditionally provided to move the piston to its rest positioncorresponding to a fully released clutch arrangement. The reinforcementof the releasing movement provided by the sealing arrangement and/orboundary wall is particularly advisable insofar as it often comes aboutthat the clutch arrangement can be released quickly, for example, whenthe clutch arrangement is to be operated with regulated slip. Variousfactors can work against a quick release, for example, inertial forceswhich are exerted on the piston by a cooling fluid, especially coolingoil, or inertial forces which are mediated by the cooling fluid and acton the piston or which are based on the mass moment of inertia; as arule, these inertial forces are particularly prominent when the pistonreleasing movement is introduced. For this reason, it is useful toreinforce the introduction of the piston releasing movement in thesuggested manner above all.

The sealing arrangement can have a sealing element which is arranged ata wall portion defining the pressure compensation chamber and can engagetightly with the latter and with the actuating piston. The sealingelement can project from the wall portion in the direction of the stopface of the actuating piston and, at least in the engaged state of theclutch arrangement, can transmit elastic restoring forces to theactuating piston in the direction of a stroke position corresponding toa released state of the clutch arrangement. The sealing element can beinjection-molded at an edge area of the wall portion or can be fittedover the edge area. This is particularly simple in terms of assemblyand, moreover, provides for a secure positioning of the sealing elementat the wall portion.

According to a second independent aspect of the invention, for adouble-clutch device, for arranging in a drivetrain of a motor vehiclebetween a drive unit and a transmission, wherein the clutch device has afirst clutch arrangement associated with a first transmission inputshaft of the transmission and a second clutch arrangement associatedwith a second transmission input shaft of the transmission fortransmitting torque between the drive unit and the transmission, it issuggested that at least one of the clutch arrangements has an actuatingpiston defining a pressure chamber for actuation, preferably forengagement of the clutch arrangement by means of a pressure medium,preferably a hydraulic medium, and that the clutch arrangements areconstructed as plate clutch arrangements, of which a radial outer clutcharrangement annularly encloses a radial inner clutch arrangement,wherein the actuating piston associated with the radial outer clutcharrangement is guided at a plate carrier, preferably an outer platecarrier, of the radial outer clutch arrangement and at a plate carrier,preferably an outer plate carrier, of the radial inner clutcharrangement so as to be axially displaceable and so as to seal thepressure chamber.

As a result of the above-mentioned guiding of the actuating piston atthe plate carriers, the actuating piston is axially guided in a securemanner so that high operating reliability is achieved and, on the otherhand, an especially compact construction of the clutch device is madepossible, since no separate guide elements are required for guiding thepiston and installation space can accordingly be economized on. When acentrifugal force pressure compensation chamber is provided, whichreceives the pressure compensation medium and is defined or limited bythe actuating piston, the actuating piston can also guide thecentrifugal force pressure compensation chamber at the plate carrier ina sealing manner.

With respect to an especially compact construction of the clutch device,it is further suggested that the centrifugal force pressure compensationchamber associated with the radial outer clutch arrangement is definedby a plate carrier, preferably an outer plate carrier, of the radialinner clutch arrangement. Accordingly, a separate wall for forming thepressure compensation chamber can be completely dispensed with, ifdesired.

Installation space can also be economized on in a corresponding mannerin that an actuating piston associated with the radial inner clutcharrangement is guided so as to be axially displaceable at a platecarrier, preferably an outer plate carrier, of the radial inner clutcharrangement and, as the case may be, at a wall of a centrifugal forcepressure compensation chamber, if provided, which is associated with theactuating piston, and so as to seal the associated pressure chamber and,as the case may be, the centrifugal force pressure compensation chamberwhich receives the pressure compensation medium. In this connection, itis advisable that sealing elements cooperating with the actuatingpistons are arranged so as to be radially staggered and preferablypartially overlapping axially.

With respect to an optimum utilization of the installation space, it canbe provided that at least one of the centrifugal force pressurecompensation chambers extends over a different radial area than theassociated pressure chamber in such a way that an effective pressureapplication surface of the piston on the pressure chamber side issmaller than an effective pressure application surface of the piston onthe pressure compensation chamber side and/or that a pressurecompensation chamber limiting surface of the piston extends fartherradially outward than a pressure chamber limiting surface of the piston.In order to avoid an overcompensation of the centrifugal force whichcould possibly occur when the pressure compensation chamber iscompletely filled with pressure compensation medium (in this regard, theradius dependency of the pressure increase in the pressure chamber andin the pressure compensation chamber caused by centrifugal force must betaken into account), fill level limiting means can be allocated to thepressure compensation chamber to limit the filling of the pressurecompensation chamber with pressure compensation medium to a maximumradial partial filling level. The fill level limiting means can compriseat least one pressure compensation medium through-opening in a wall ofthe pressure compensation chamber extending in radial direction.

According to a third independent aspect of the invention, for adouble-clutch device, for arranging in a drivetrain of a motor vehiclebetween a drive unit and a transmission, which clutch device has a firstclutch arrangement associated with a first transmission input shaft ofthe transmission and a second clutch arrangement associated with asecond transmission input shaft of the transmission for transmittingtorque between the drive unit and the transmission, it is suggested thatat least one of the clutch arrangements has an actuating piston defininga pressure chamber for actuation, preferably for engagement of theclutch arrangement by means of a pressure medium, preferably a hydraulicmedium, wherein the actuating piston is guided at a wall portiondefining the pressure chamber and/or at a wall portion defining anassociated pressure compensation chamber and/or so as to seal thepressure compensation chamber, wherein a labyrinth seal comprising atleast one annular groove in a surface portion of the actuating pistonand/or of the wall portion, which surface portion extends in anengaging-disengaging direction of the actuating piston, acts between therespective wall portion and the piston, and/or wherein at least onesealing ring acts between the respective wall portion and the piston,which sealing ring is secured to one actuating piston and wall portionon the one hand and acts at the other actuating piston and wall portionwith axial play relative to one actuating piston and wall portion on theother hand and is acted upon during an engaging movement and/or during adisengaging movement of the actuating piston such that the sealingengagement with the wall portion and/or actuating piston increases.

When a labyrinth seal of the type mentioned above is provided, separatesealing elements made of rubber, plastic or elastomer material, forexample, can be dispensed with and the seal and associatedcounter-element can be made of the same material with, consequently, thesame thermal coefficient of expansion. It is extremely advantageous forboth sealing members of the labyrinth seal to have the same thermalcoefficient of expansion insofar as temperature changes or temperaturefluctuations cause no changes, or no substantial changes, in theengagement or frictional engagement, as the case may be, between thesealing members which would lead to corresponding fluctuations ininertial forces to be overcome during the engaging and disengagingmovement of the piston. Further, there are no substantial fluctuationsin the sealing action, especially no loss of tightness.

The other constructional variant according to the invention with thesealing ring which is fixed on one side and engages with axial play onthe other side is advantageous insofar as particularly good sealing canbe achieved, for example, when the sealing ring is additionally actedupon by the pressure in the pressure chamber or pressure compensationchamber to strengthen the sealing engagement in such a way that thesealing ring is pressed more firmly into a shaped portion receiving thesealing ring. This takes place to a marked degree in the case of thepressure chamber when the clutch is engaged with the stationary pistonin its moved out or disengaged position, i.e., when the plate stack istensioned to a maximum degree and as little leakage as possible shouldoccur.

It is assumed in the preceding remarks that the sealing ring is actedupon to increase the sealing engagement during an engaging movement andreaches a maximum sealing engagement in the end position of theactuating piston. However, if it should seem advisable, the maximumsealing engagement could also be provided in a middle stroke position ofthe piston or in the end position of the actuating piston correspondingto a released clutch.

The sealing ring can engage in an annular groove of the actuating pistonor of the wall portion. Further, the sealing ring can have at least oneportion which acts, for example, in the manner of a stripper, at asurface region of the actuating piston or of the wall portion whichextends in an engaging-disengaging direction of the actuating piston andwhich is plane in this direction.

It is suggested that the sealing ring is curved in axial direction in astate of lower tension with a looser sealing engagement with respect toa section plane containing or parallel to an axis of rotation of theclutch device and is moved to a state of higher tension with a tightersealing engagement during an engaging movement and/or during adisengaging movement of the actuating piston in that it is stretchedwith respect to the section plane or is less sharply curved in axialdirection.

As was already indicated, it is preferred in general that the sealingring is acted upon to the maximum sealing engagement in the engagedstate of the respective clutch arrangement.

The wall portion defining the pressure chamber can be a plate carrier,preferably an outer plate carrier, of the respective clutch arrangementconstructed as plate clutch arrangement.

When the sealing ring is inserted into a shaped portion or cutout of theactuating piston on one side and acts at the plane surface portion onthe other side, axial installation space can be economized on in thatthe shaped portion is provided in the edge area of a radially extendingpiston portion. Thin wall thicknesses are made possible in this way. Theshaped portion, possibly, a groove, can then be rolled in if desired. Inorder to economize on axial installation space, it is also helpful thatthe sealing ring can be elongated in cross section in radial directionas suggested according to the invention without the risk of impairingthe sealing action.

According to a fourth aspect of the invention, for a double-clutchdevice, for arranging in a drivetrain of a motor vehicle between a driveunit and a transmission, which clutch device has a first clutcharrangement associated with a first transmission input shaft of thetransmission and a second clutch arrangement associated with a secondtransmission input shaft of the transmission for transmitting torquebetween the drive unit and the transmission, it is suggested that eachof the clutch arrangements has an actuating piston defining a pressurechamber for actuation, preferably for engagement of the clutcharrangement by means of a pressure medium, preferably a hydraulicmedium, wherein a radial outer sealing element (designated as firstsealing element with respect to the pressure chamber) which seals thepressure chamber of the first clutch arrangement on the radial outerside and/or axially and which acts between the actuating piston and awall of the pressure chamber and a radial outer sealing element(designated as second sealing element with respect to the pressurechamber) which seals the pressure chamber of the second clutcharrangement on the radial outer side and/or axially and which actsbetween the actuating piston and a wall of the pressure chamber arearranged at different radial distances from an axis of rotation of theclutch device, and/or wherein a radial inner sealing element (designatedas third sealing element with respect to the pressure chamber) whichseals the pressure chamber of the first clutch arrangement at the radialinner side and/or axially and which acts between the actuating pistonand a wall of the pressure chamber and a radial inner sealing element(designated as fourth sealing element with respect to the pressurechamber) which seals the pressure chamber of the second clutcharrangement at the radial inner side and/or axially and which actsbetween the actuating piston and a wall of the pressure chamber arearranged at different radial distances from an axis of rotation of theclutch device.

Due to the different radial spacing of the sealing elements providedaccording to the invention, the radial and/or axial installation spaceof the clutch device can be kept comparatively small. In order to makepossible a particularly compact construction of the clutch device, it issuggested in a further development that the first and second sealingelements and/or the third and fourth sealing elements are so arrangedwith respect to a section plane containing or parallel to the axis ofrotation of the clutch device that a straight line which intersects bothsealing elements encloses, with an axis of rotation of the clutchdevice, an angle α of approximately 10° to 70°, preferably approximately20° to 50°, most preferably approximately 30° to 40°, at least in thedisengaged state of the two clutch arrangements and/or in the engagedstate of the two clutch arrangements.

The actuating piston preferably divides the respective pressure chamberfrom an associated pressure compensation chamber which receives apressure compensation medium. For this purpose, it is suggested by wayof a further development that a radial outer sealing element (designatedas fifth sealing element with respect to the pressure compensationchamber) which seals the pressure compensation chamber of the firstclutch arrangement on the radial outer side and/or axially and whichacts between the actuating piston and a wall of the pressurecompensation chamber and a radial outer sealing element (designated assixth sealing element with respect to the pressure compensation chamber)which seals the pressure compensation chamber of the second clutcharrangement on the radial outer side and/or axially and which actsbetween the actuating piston and a wall of the pressure compensationchamber are arranged at different radial distances from an axis ofrotation of the clutch device, and/or that a radial inner sealingelement (designated as seventh sealing element with respect to thepressure compensation chamber) which seals the pressure compensationchamber of the first clutch arrangement on the radial inner side and/oraxially and which acts between the actuating piston and a wall of thepressure compensation chamber and a radial inner sealing element(designated as eighth sealing element with respect to the pressurecompensation chamber) which seals the pressure compensation chamber ofthe second clutch arrangement on the radial inner side and/or axiallyand which acts between the actuating piston and a wall of the pressurecompensation chamber are arranged at different radial distances from anaxis of rotation of the clutch device.

According to this suggestion, a construction of the clutch device whichis comparatively compact radially and/or axially can be achieved inspite of the provided pressure compensation chambers especially when thefifth and sixth sealing elements and/or the seventh and eighth sealingelements are so arranged with respect to a section plane containing orparallel to the axis of rotation of the clutch device that a straightline which intersects both sealing elements encloses, with an axis ofrotation of the clutch device, an angle α of approximately 10° to 70°,preferably approximately 30° to 60°, most preferably approximately 40°to 50°, at least in the disengaged state of the two clutch arrangementsand/or in the engaged state of the two clutch arrangements.

It is suggested for the sealing elements mentioned above that the firstand the fifth sealing element are separate sealing elements which arepreferably arranged at an essentially identical radial distance from theaxis of rotation of the clutch device, and/or that the second and thesixth sealing elements are separate sealing elements which arepreferably arranged at an essentially identical radial distance from theaxis of rotation of the clutch device, and/or that the third and theseventh sealing elements are formed by a sealing element associated withboth the pressure chamber and the pressure compensation chamber, and/orthat the fourth and the eighth sealing elements are formed by a sealingelement associated with both the pressure chamber and the pressurecompensation chamber. In general, however, the seals can also bearranged on different diameters, wherein a matching of the pressures inthe piston chambers can be achieved by selecting the diameters so as tobe adapted to one another.

The features of a multiple-clutch device and a drivetrain which wereindicated in connection with the different aspects of the invention canbe advantageously combined. Further independent aspects of the inventionwill be discerned by the person skilled in the art from the precedingexplanations and the description of the Figures.

The invention is further directed to a drivetrain for a motor vehiclewith a clutch device, according to at least one aspect of the invention,arranged between a drive unit and a transmission.

The invention will be described more fully in the following withreference to embodiment examples shown in the Figures.

The various features of novelty which characterize the invention arepointed out with particularity in the claims annexed to and forming apart of the disclosure. For a better understanding of the invention, itsoperating advantages, and specific objects attained by its use,reference should be had to the drawing and descriptive matter in whichthere are illustrated and described preferred embodiments of theinvention.

BRIEF DESCRIPTION OF THE DRAWINGS

In the drawings:

FIG. 1 is a view in partial section showing a double-clutch with twoplate clutch arrangements which is arranged in a drivetrain of a motorvehicle between a transmission and a drive unit;

FIG. 2 is a view similar to FIG. 1 depicting a modified seal with whichthe lid is sealed to the clutch housing.

FIG. 3 is a view similar to FIG. 1 showing another form of seal of thelid to the clutch housing.

FIG. 4 is a view of a variant of the double clutch construction of FIG.3.

FIG. 5 depicts an advantageous manner of arranging sealing elements inthe double clutch to provide a more compact contraction.

FIG. 6 is a view depicting an oil drain channel provided in the clutchhousing.

FIGS. 7a) and 7 b) are showings of sealing the lid with the clutchhousing wherein a sealing ring is carried in the lid and in the housingrespectively.

FIGS. 7c) and 7 d) show alternate ways of sealing the pressure chambersand the pressure compensation chambers in the dowlde clutch.

FIG. 8 depicts sealing of the lid with the clutch housing with anannular ring element of rubber or plastic as well as use of an annularsecuring plate for axially securing of the ring element.

FIG. 9 shows sealing of the clutch space with a sealing compound.

FIG. 10 a further manner of sealing the clutch space using sealingcompound.

FIGS. 11 and 12 show further aspects of the clutch construction.

FIG. 13 illustrates how the clutch device is coupled to the drivetrainvia a clutch hub, that is, preferably via a torsional vibration damper.

FIG. 14 shows a hub sealing arrangement.

DETAILED DESCRIPTION OF THE PRESENTLY PREFERRED EMBODIMENTS

FIG. 1 shows a double-clutch 12 arranged in a drivetrain 10 between adrive unit and a transmission. The drive unit, for example, an internalcombustion engine, is represented in FIG. 1 only by a driven shaft 14,such as a crankshaft 14, with a coupling end 16 serving to connect atorsional vibration damper, not shown. In FIG. 1, the transmission isrepresented by a transmission housing portion 20 defining a transmissionhousing cover 18 and by two transmission input shafts 22 and 24, both ofwhich are constructed as hollow shafts, wherein transmission input shaft22 extends through transmission input shaft 24 substantially coaxialthereto. A pump drive shaft which serves to drive a transmission-sideoil pump, not shown in FIG. 1, as will be described more fully in thefollowing, is arranged in the interior of the transmission input shaft22.

The double-clutch 12 is received in the transmission housing cover 18,wherein the interior of the cover is closed in the direction of thedrive unit by a lid 28 which is pressed into a cover housing openingand/or is secured therein by a snap ring 30. When the double-clutch haswet-type friction clutches as in the embodiment example shown in FIG. 1,for example, diaphragm clutches, it is generally arranged to provide fora tight engagement between the lid 28 and the transmission housingformed by the transmission housing cover 18, which can be produced, forexample, by an O-ring or other sealing ring. FIG. 1 shows a sealing ring32 with two sealing lips.

A clutch hub 34 which is formed of two ring portions 36, 38 secured toone another for reasons which will be explained more fully serves as theinput side of the double-clutch 12. The clutch hub 34 extends through acentral opening of the lid 28 in the direction of the drive unit and iscoupled with the torsional vibration damper, not shown, via an externaltoothing 42, so that there is a torque-transmission connection betweenthe coupling end 16 of the crankshaft 14 and the clutch hub 34 by way ofthis torsional vibration damper. If it is desirable to dispense with atorsional vibration damper in general, or at this location in thedrivetrain, the clutch hub 34 can also be coupled directly with thecoupling end 16. At its end remote of the transmission, the pump driveshaft 26 has an external toothing 44 which engages in an internaltoothing 46 of the ring portion 36 of the clutch hub 34, so that thepump drive shaft 26 rotates along with the clutch hub 34 andconsequently drives the oil pump when a rotational movement is impartedto the clutch hub 34, as a rule, by the drive unit and in many operatingsituations possibly also by the transmission via the double-clutch (forexample, in the operating situation characterized under the heading of“engine brake”).

The lid 28 extends radially between an annular circumferential wallportion of the housing cover 18, which defines a radial recess 50 of thehousing cover 18, and the ring portion 38 of the hub 34. It isadvantageous when a sealing arrangement and/or pivot bearing arrangement54 are/is provided between a radial inner wall area 52 of the lid 28 andthe hub 34, especially the ring portion 38, particularly when—as in theshown embodiment example—the lid 28 is secured to the housing cover 18and consequently does not rotate along with the double-clutch 12.Sealing between the lid and the hub is required particularly when theclutch arrangements of the double-clutch are wet-type clutches as in theembodiment example. A high degree of operational reliability is alsoachieved in case of oscillations and vibrations when the sealing and/orpivot bearing arrangement 54 is secured axially to the lid 28 and/or tothe clutch hub 34, for example, by means of an end portion of the lidedge 52 which is bent radially inward as is shown in FIG. 1.

A carrier plate 60 is arranged at the ring portion 38 of the hub 34 soas to be fixed with respect to rotation relative to it and serves totransmit torque between the hub 34 and an outer plate carrier 62 of afirst plate clutch arrangement 64. The outer plate carrier 62 extends inthe direction of the transmission and radially inward to a ring part 66at which the outer plate carrier is arranged so as to be fixed withrespect to rotation relative to it, this ring part 66 being supported atthe two transmission input shafts 22 and 24 by means of an axial andradial bearing arrangement 68 in such a way that both radial and axialforces are supported at the transmission input shafts. The axial andradial bearing arrangement 68 enables relative rotation between the ringpart 66 on the one hand and the transmission input shaft 22 andtransmission input shaft 24 on the other hand. The construction andoperation of the axial and radial bearing arrangement will be discussedat greater length in the following.

An outer plate carrier 70 of a second plate clutch arrangement 72 isarranged at the ring part 66 farther axially in the direction of thedrive unit so as to be fixed with respect to rotation relative to it,the plate stack 74 of the second plate clutch arrangement 72 beingenclosed annularly by plate stack 76 of the first plate clutcharrangement. As was already indicated, the two outer plate carriers 62and 70 are connected with one another by the ring part 66 so as to befixed with respect to relative rotation and, together, by way of thecarrier plate 60 which is in a positive-locking torque-transmittingengagement with the outer plate carrier 62 by means of an externaltoothing, are in a torque-transmitting connection with the clutch hub 34and therefore—via the torsional vibration damper, not shown—with thecrankshaft 14 of the drive unit. With respect to the normal flow oftorque from the drive unit to the transmission, the outer plate carriers62 and 70 serve as the input side of the plate clutch arrangement 64 and72, respectively.

A hub part 80 of an inner plate carrier 82 of the first plate clutcharrangement 64 is arranged on the transmission input shaft 22 so as tobe fixed with respect to rotation relative to it by means of a keywaytoothing or the like. In corresponding manner, a hub part 84 of an innerplate carrier 86 of the second plate clutch arrangement 72 is arrangedon the radially outer transmission input shaft 24 so as to be fixed withrespect to rotation relative to it by means of a keyway toothing or thelike. With respect to the normal torque flow from the drive unit in thedirection of the transmission, the inner plate carriers 82 and 86 serveas output side of the first and second plate clutch arrangement 64 and72, respectively.

Referring again to the radial and axial bearing support of the ring part66 at the transmission input shafts 22 and 24, two radial bearingsubassemblies 90 and 92 which act between the radial outer transmissioninput shaft 24 and the ring part 66 serve to support the ring part 66radially. With respect to support in the direction of the drive unit,the axial bearing support of the ring part 66 is carried out by hub part84, axial bearing 94, hub part 80 and a snap ring 96 which secures thehub part 80 axially to the radial inner transmission input shaft 22. Thering part 38 of the clutch hub 34 is again supported via an axialbearing 98 and a radial bearing 100 at the hub part 80. The hub part 30is supported axially in the direction of the transmission via the axialbearing 94 at an end portion of the radial outer transmission inputshaft 24. The hub part 84 can be supported directly at an annular stopor the like or at a separate snap ring or the like in the direction ofthe transmission at the transmission input shaft 24. Since the hub part84 and the ring part 66 are rotatable relative to one another, an axialbearing can be provided between these components insofar as the bearing92 does not perform both the function of an axial bearing and thefunction of a radial bearing. The latter is assumed in the embodimentexample in FIG. 1.

As is shown in the embodiment example, it is very advantageous when theportions of the outer plate carriers 62 and 70 extending in radialdirection are arranged on an axial side of a radial plane extending toan axis A of the double-clutch 12 and the portions of the inner platecarriers 82 and 86 of the two plate clutch arrangements extending inradial direction are arranged on the other axial side of this radialplane. In this way, a particularly compact construction is possible,especially when—as is shown in the embodiment example—plate carriers ofone type (outer plate carrier, as in the embodiment example, or innerplate carrier) are connected with one another so as to be fixed withrespect to rotation and serve in each instance as the input side of therespective plate clutch arrangement with respect to the flow of forcefrom the drive unit to the transmission.

Actuating pistons are integrated in the double-clutch 12 for actuatingthe plate clutch arrangements, in the case of the embodiment exampleshown, for actuating the plate clutch arrangements for the purpose ofengagement. An actuating piston 110 associated with the first plateclutch arrangement 64 is arranged axially between the radially extendingportion of the outer plate carrier 62 of the first plate clutcharrangement 64 and the radially extending portion of the outer platecarrier 70 of the second plate clutch arrangement 72 and is axiallydisplaceable at both outer plate carriers and at the ring part 66 bymeans of seals 112, 114, 116 and guided so as to seal a pressure chamber118 formed between the outer plate carrier 62 and the actuating piston110 and a centrifugal force pressure compensation chamber 120 formedbetween the actuating piston 110 and the outer plate carrier 70. Thepressure chamber 118 communicates, via a pressure medium channel 122formed in the ring part 66, with a pressure control device, e.g., acontrol valve, connected to a pressure medium supply, in this case, theoil pump mentioned above. The pressure medium channel 122 is connectedto the pressure control device via a connection sleeve which receivesthe ring part 66 and which is fixed relative to the transmission. Inthis connection, it should be added with reference to the ring part 66that this ring part 66 is produced in two parts with two sleeve-likering part portions inserted one inside the other as is indicated in FIG.1 in order to simplify production especially with regard to the pressuremedium channel 122 and an additional pressure medium channel.

An actuating piston 130 associated with the second plate clutcharrangement 72 is arranged axially between the outer plate carrier 70 ofthe second plate clutch arrangement 72 and a substantially radiallyextending wall portion 132, which is arranged at an axial end area ofthe ring part 66 remote of the transmission so as to be fixed withrespect to relative rotation it and so as to be tight against fluid, andis guided so as to be axially displaceable by means of seals 134, 136and 138 at the outer plate carrier 70, wall portion 132 and ring part 66and is guided so as to seal a pressure chamber 140 formed between theouter plate carrier 70 and the actuating piston 130 and a centrifugalforce pressure compensation chamber 142 formed between the actuatingpiston 130 and the wall portion 132. The pressure chamber 140 isconnected via another pressure medium channel 144 (mentioned above) in amanner corresponding to pressure chamber 118 at a/the pressure controldevice. Pressure can be applied by means of the pressure controldevice(s) to the two pressure chambers 118 and 140 selectively (possiblyalso simultaneously) from the pressure medium source (in this case, theoil pump) in order to actuate the first plate clutch arrangement 64and/or the second plate clutch arrangement 72 for purposes ofengagement. Diaphragm springs 146, 148, are used for resetting, that is,for releasing the clutches; the diaphragm spring 148 associated with theactuating piston 130 is received in the centrifugal force pressurecompensation chamber 142.

The pressure chambers 118 and 140 are completely filled with pressuremedium (in this case, hydraulic oil) in every case during normaloperating states of the double-clutch 12, and the actuating state of theplate clutch arrangements depends on the pressure of the pressure mediumPresent at the pressure chambers. However, since the outer platecarriers 62 and 70, including the ring part 66, the actuating pistons110 and 130 and the wall portion 132, rotate along with the crank shaft14 during driving operation, pressure increases due to centrifugal forceoccur in the pressure chambers even when no pressure is applied to thepressure chambers 118 and 140 proceeding from the pressure controldevice; these increases in pressure could lead to unwanted engagement orat least grinding of the plate clutch arrangements at least at higherrotational speeds. The centrifugal force pressure compensation chambers120, 142 mentioned above are provided for this reason. These centrifugalforce pressure compensation chambers 120, 142 hold a pressurecompensation medium, and pressure increases caused by centrifugal forceare brought about therein in a corresponding manner to compensate forthe pressure increases due to centrifugal force which occur in thepressure chambers.

It is possible to fill the centrifugal force pressure compensationchambers 120 and 142 permanently with pressure compensation medium, forexample, oil, wherein volume compensation could be provided, if need be,for receiving pressure compensation medium which is displaced in thecourse of actuating the actuating pistons. In the embodiment form shownin FIG. 1, the centrifugal force pressure compensation chambers 120, 142are first filled with pressure compensation medium when the drivetrainis in operation, namely, in connection with the supply of cooling fluid,especially cooling oil as in the embodiment example, to the plate clutcharrangements 64 and 72 via an annular channel 150 formed between thering part 66 and the outer transmission input shaft 24, wherein thebearings 90, 92 through which cooling oil can pass are associated withthis ring channel 150. The cooling oil flows from a transmission-sideconnection between the ring part and transmission input shaft 24 in thedirection of the drive unit through bearing 90 and bearing 92 and thenflows in a partial flow between the end portion of the ring part 66remote of the transmission and the hub part 84 radially outward in thedirection of the plate stack 74 of the second plate clutch arrangement72, enters the area of the plates because of the through-openings in theinner plate carrier 86, flows between the plates of the plate stack 74and radially outward through friction facing grooves or the like inthese plates, enters the area of plate stack 76 of the first plateclutch arrangement 64 through through-openings in the outer platecarrier 70 and through-openings in the inner plate carrier 82, flowsradially outward between the plates of this plate stack or thoughfriction facing grooves or the like in these plates, and then finallyflows radially outward through through-openings in the outer platecarrier 62. The centrifugal force pressure compensation chambers 120,142 are also connected to the cooling oil feed flow between the ringpart 66 and the transmission input shaft 24, namely, by means of radialbore holes 152, 154 in the ring part 66. Since the cooling oil servingas pressure compensation medium in the pressure compensation chambers120, 142 runs out of the pressure compensation chambers due to theabsence of centrifugal forces when the drive unit is stationary, each ofthe pressure compensation chambers is re-filled during operation of thedrivetrain (of the motor vehicle).

Since a pressure application surface of the actuating piston 130associated with the pressure chamber 140 is smaller and, moreover,extends less far radially outward than a pressure application surface ofthe piston 130 associated with the pressure compensation chamber 142, atleast one fill level limiting opening 156 which adjusts a maximum radialfilling level of the pressure compensation chamber 142 giving therequired centrifugal force compensation is formed in the wall portion132. When the maximum filling state is reached, the cooling oil suppliedvia the bore hole 154 flows through the filling level limiting opening156 and unites with the cooling oil flow passing radially outwardbetween the ring part 66 and hub part 84. With respect to the piston110, the pressure application surfaces of the piston associated with thepressure chamber 118 and the pressure compensation chamber 120 are thesame size and extend within the same radial area, so that correspondingfill level limiting means are not required for the pressure compensationchamber 120.

For the sake of completeness, it should be mentioned that additionalcooling flows preferably occur in operation. Accordingly, at least oneradial bore hole 160 is provided in the transmission input shaft 24,wherein another cooling oil partial flow flows through this radial borehole 160 and through an annular channel between the two transmissioninput shafts. This other cooling oil partial flow divides into twopartial flows, one of which flows radially outward between the two hubparts 80 and 84 (through the axial bearing 94) and the other partialflow flows radially outward between the end area of the transmissioninput shaft 22 remote of the transmission and the hub part 80 andbetween this hub part 84 and the ring portion 38 of the clutch hub 34(through the bearings 98 and 100).

Since the cooling oil flowing radially outward could accumulate next toa radially outer portion of the actuating piston 110 associated with thefirst plate clutch arrangement 64 and could impede the engaging movementof this piston due to centrifugal force at least at higher rotationalspeeds, the piston 110 has at least one pressure compensation opening162 which enables a cooling oil flow from one side of the piston to theother. Consequently, an accumulation of cooling oil will come about onboth sides of the piston with corresponding compensation of pressureforces exerted on the piston due to centrifugal force. Further, otherforces based on an interaction of the cooling oil with the piston areprevented from impeding the required axial piston movements. Thisrefers, for example, to hydrodynamic forces or the like and suctionattachment of the piston to the outer plate carrier 62.

It is also possible to provide at least one cooling oil outlet openingin the radially extending, radially outer area of the outer platecarrier 62 of the first plate clutch arrangement 64. A cooling oiloutlet opening of this kind is indicated in dashes at 164. In order toensure a sufficient flow of cooling fluid (cooling oil) through theplate stack 76 of the first plate clutch arrangement 64 in spite ofthis, a cooling oil conducting element (generally, a cooling fluidconducting element) can be provided. It is indicated in dashed lines inFIG. 1 that an adjacent end plate 166 of the plate stack 76 could have acooling oil conducting portion 168, so that the end plate 166 itselfserves as a cooling oil conducting element.

With respect to a simple construction of the pressure control device forthe actuation of the two plate clutch arrangements, it was provided inthe embodiment example of FIG. 1 that a torque transmitting capacitywhich is given, per se, for the radial inner plate clutch arrangement 72with reference to an actuating pressure and which is smaller compared tothe other clutch arrangement 64 (because of a smaller effective frictionradius than the radial outer clutch arrangement 64) is at leastpartially compensated. For this purpose, the pressure applicationsurface of the piston 130 associated with the pressure chamber 140 islarger than the pressure application surface of the piston 110associated with the pressure chamber 118, so that axially directedforces greater than those exerted on the piston 110 are exerted onpiston 130, given the same hydraulic oil pressure in the pressurechambers.

It should be mentioned that the available installation space is made useof in a favorable manner due to a radial staggering of the sealsassociated with the piston, especially also an axial overlapping of atleast some of the seals.

In addition to the above-mentioned supply of cooling oil and the formingof cooling oil through-openings (indicated only schematically in FIG. 1)in the plate carriers, steps can be taken in the plate stacks 74, 76 toprevent the risk of overheating. Accordingly, at least some of theplates are advantageously used as “heat buffers” that temporarily storeheat which is formed, for example, during slip operation and whichtemporarily overburdens the heat dissipation possibilities allowed bycooling fluid (in this case, cooling oil) or by heat conductance via theplate carriers; in this way, the heat can be carried off at a latertime, for instance, in a disengaged state of the respective plate clutcharrangement. For this purpose, the plates in the radial inner (second)plate clutch arrangement which have no friction facings are constructedso as to be thicker axially than plate carrier elements of plates havingfriction facings in order to provide a comparatively large materialvolume with corresponding heat capacity for plates without frictionfacings. These plates should be made of a material having a considerableheat storage capability (heat capacity), for example, steel. The plateshaving friction facings can temporarily store only a little heat whenusing conventional friction facings made from paper, for example, sincepaper has poor heat conductivity.

The heat capacity of the elements having friction facings can likewisebe made available as heat storage when facing materials with highconductivity are used instead of facing materials with low conductivity.It is possible to use friction facings of sintered material which has acomparatively high heat conductivity. However, the problem with the useof sintered facings is that the sintered facings have a degressive curveof the coefficient of friction μ over slippage speed (relativerotational speed ΔN between the rubbing surfaces); that is, dμ/dΔb N <0.A degressive curve of the coefficient of friction is disadvantageousinsofar as it can promote self-excitation of oscillations in thedrivetrain or, at least, cannot damp such oscillations. Therefore, it isadvantageous when plates with friction facings of sintered material aswell as plates with friction facings of another material with aprogressive curve of the coefficient of friction over the slippage speed(dμ/dΔN>0) are provided in a plate stack, so that a progressive curve ofthe coefficient of friction over the slippage speed results for theplate stack as a whole or there is at least approximately a neutralcurve of the coefficient of friction over the slippage speed (dμ/dΔN=0)and, consequently, self-excitation of oscillations in the drivetrain isat least not promoted or, preferably, torsional vibrations in thedrivetrain are even damped (because of a considerable progressive curveof the coefficient of friction over the slippage speed).

It is assumed in this connection that in the embodiment example of FIG.1 the plate stack 74 of the radial inner plate clutch arrangement 60 isconstructed without sintered facings, since the radial outer plateclutch arrangement 64 is preferably used as a starting clutch withcorresponding slip operation. The latter, that is, the use of the radialouter plate clutch arrangement as a starting clutch, is advantageousinsofar as this plate clutch arrangement can be operated with loweractuating forces (for the same torque transmitting capacity) because ofthe larger effective friction radius, so that the area pressure can bereduced relative to the second plate clutch arrangement. For thispurpose, it is also helpful when the plates of the first plate clutcharrangement 64 are formed with a somewhat greater radial height than theplates of the second plate clutch arrangement 72. However, if desired,friction facings of sintered material can also be used for the platestack 74 of the radial inner (second) plate clutch arrangement 72,preferably, as was already mentioned, in combination with frictionfacings of another material such as paper.

In the plate stack 74 of the radial inner plate clutch arrangement 72,all inner plates have friction facings and all outer plates are withoutfriction facings, wherein the end plates defining the plate stackaxially are outer plates and accordingly have no facings; in the platestack 76 of the first plate clutch arrangement 64, the inner plates haveno facings and the outer plates, including the end plates 166, 170, havefriction facings. According to a preferred construction, at least theend plates 166 and 170 have facing-carrying elements which aresubstantially thicker axially than the facing-carrying elements of theother outer plates and are formed with facings of sintered material, sothat the facing-carrying elements of the two end plates which have acomparatively large volume can be put to use as heat buffers. As withthe plate stack 74, the plates having no facings are thicker axiallythan the plate carrying elements of the plates having friction facings(with the exception of the end plates) in order to provide acomparatively large heat capacity for temporary storage of heat. Theouter plates located axially inside should, at least in part, havefriction facings of a different material exhibiting a progressive curveof the coefficient of friction in order to achieve at least anapproximately neutral curve of the coefficient of friction over theslippage speed for the plate stack as a whole.

Further details of the double-clutch 12 according to the describedembodiment example can be readily discerned from FIG. 1 by the personskilled in the art. For example, the axial bore hole in the ring portion36 of the clutch hub 34 in which is formed the internal toothing 46 forthe pump drive shaft is closed so as to be tight against oil by means ofa stopper 180 secured therein. The carrier plate 60 is fixed axially tothe outer plate carrier 62 by two retaining rings 172, 174, whereinretaining ring 172 also supports the end plate 170 axially. Acorresponding retaining ring is also provided for supporting the platestack 74 at the outer plate carrier 70.

With respect to the construction of the outer plates of the first plateclutch arrangement 64 as facing-carrying plates, it should be added thatan improved through-flow through the plate stack 76 is achieved inconnection with the allocation of the outer plates to the input side ofthe clutch device when the friction facings—as is generally the case—areformed with friction facing grooves or other fluid passages which enableflow through the plate stack also in the state of frictional engagement.Since the input side also rotates along with the running drive unit andthe coupling end 16 when the clutch arrangement is released, a kind ofconveying action is brought about because of the revolving frictionfacing grooves and the revolving fluid passages, so that thethrough-flow through the plate stack is improved in a correspondingmanner. In contrast to the view in FIG. 1, the second plate clutcharrangement could also be constructed accordingly, that is, the outerplates could be constructed as plates having friction facings.

Referring to FIGS. 2 to 14, further embodiment examples of themultiple-clutch devices according to the invention, especiallydouble-clutch devices according to the invention, will now be describedin relation to various aspects. Since the embodiment examples of FIGS. 2to 14 correspond to the embodiment example of FIG. 1 with respect tobasic construction and the views shown in FIGS. 2 to 14 will beimmediately understood by the person skilled in the art based on thepreceding detailed explanation of the embodiment example of FIG. 1, itwill not be necessary to describe the embodiment examples in FIGS. 2 to14 in all particulars. In this connection, reference is had to thepreceding explanation of the embodiment example of FIG. 1 which can becarried over to a great extent to the embodiment examples in FIGS. 2 to14. The reference numbers used for the embodiment examples of FIGS. 2 to14 are the same as those used for the embodiment example in FIG. 1. Forthe sake of clarity, not all of the reference numbers in FIG. 1 are alsoshown in FIGS. 2 to 14 insofar as the double-clutches of the embodimentexamples in FIGS. 2 to 14 correspond to the embodiment example of FIG.1.

An important aspect for clutch devices with wet-type clutch arrangementsis the sealing of the clutch space and, in connection with this, thefixing of the lid 28 in the opening of the clutch housing 20. In theembodiment examples in FIGS. 3, 6 and 7, the lid 28 is over dimensionedradially and is pressed into the opening of the clutch housing formed bythe housing portion 20. The sealing ring 32 which seals the clutchhousing is provided because the lid can sag or become wavy under certaincircumstances. The sealing ring has the further object of damping anyoscillations with axial relative movements between the lid 28 and theclutch housing. The sealing ring, which can be constructed as an O-ring,can be mounted at the lid and/or at the housing and, for this purpose,can be received in an annular groove of the housing (compare FIG. 7b)and/or in an annular groove of the lid (compare FIG. 7a) formed in anedge portion of the lid 28. For an enhanced sealing action, two or moreO-rings which are arranged axially adjacent to one another could also beprovided instead of one O-ring. Another possibility is to use a sealingring with two or more sealing lips (compare FIG. 1 and FIG. 14).

For stricter requirements regarding tightness, the solutions applied inthe embodiment examples of FIGS. 2, 6, 8, 9, 10, 11 and 12 can beconsidered. In some of these embodiment examples (compare, e.g., FIGS. 2and 11), a rubber ring or plastic ring was inserted prior to mountingthe lid 28 or, alternatively, an annular ring element wasinjection-molded. The respective sealing element provided in this manneris designated by 200 in the Figures. This elastic element, that is, therubber or plastic ring or the injection-molded sealing element, isclamped axially between the lid 28 and the housing 20 when the lid ismounted. A double sealing is achieved in combination with the sealingring 32. Also, the sealing element 32 can often be dispensed withbecause a very good sealing action is achieved by means of the axiallyclamped sealing element. Similar to the embodiment example in FIG. 1, asnap ring 30 takes over the function of axial securing when the clampingforces possibly acting between the lid 28 and the housing 20 are notsufficient. An alternative to the snap ring is realized in theembodiment example in FIG. 5. In this case, instead of the snap ring, anannular securing plate 210 is provided which is fastened to the clutchhousing 20, e.g., by means of screws 212. Instead of an annular securingplate 210, a plurality of separate securing plate segments could also bein provided. The lid 28 is secured in this way also in the embodimentexample in FIG. 8. Instead of an annular securing plate or a pluralityof securing plate segments, screws which are screwed into the clutchhousing and which have screw heads projecting in the radial area of thelid 28 or washer elements (for example, washers or springs) could alsobe provided.

An excellent sealing of the clutch space is achieved by means of thesolutions realized in the embodiment examples in FIGS. 9 and 10. Inthese embodiment examples, a sealing compound 205, for example, asealing foam 205 (or alternatively an elastomer or the like), wassprayed on the sealing joint between the lid 28 and the housing 20 aftermounting the lid 28. This foam 205 (or, generally, this sealing compound205) can additionally take over the function of axially securing the lid28 (the snap ring 30 of the embodiment example in FIG. 9 can accordinglypossibly be dispensed with). Further, the foam 205 can dampenoscillations with axial relative movements and/or radial relativemovements between the lid 28 and the housing 20.

In order to bring any residual leakage under control, for example, whenit is desirable to make do with a particularly simple sealingarrangement, e.g., only one O-ring, an oil drain formed by a channel 220can be provided in the clutch housing 20 corresponding to the embodimentexample in FIG. 6. It is sufficient when the channel 220 is providedonly in a lower area of the clutch housing; that is, it need not beformed all around. The channel 220 can be connected with a collectingreservoir. In some cases, it is also sufficient when the channel is onlyemptied via a drain at intervals in the course of regular maintenance.

As regards a wet clutch arrangement or wet clutch arrangements, anotherlocation to be sealed is located on the radial inside between the inputside (hub 34) of the clutch device and the lid 28. Since the lid 28 isstationary and the hub 34 rotates when the drive unit is running, acorrespondingly effective sealing arrangement 54 which withstands therotation of the hub 34 relative to the lid 28 without excessive wearshould be provided; in addition, this sealing arrangement 54 canfunction as a bearing under certain circumstances. Similar to theembodiment example in FIG. 1, the sealing arrangement 54 is securedaxially in the embodiment examples of FIGS. 3, 9 and 14 by means of abent lid edge portion or “overhang” (FIG. 3, FIG. 14) or pressedmaterial at the lid edge (FIG. 9). The lid 28 can be slit in the area ofthe overhang. Otherwise, the part of the lid in the radial area of thesealing arrangement 54 should at least be closed in order to preventleaks as far as possible.

An important aspect is the bearing support of the clutch device in thedrivetrain. The clutch device is preferably supported axially andradially at the transmission input shafts 22 ad and 24 and not at allor, at most, only secondarily (for example, with the intermediary of thelid 28 and/or a connection sleeve receiving the ring part 66) at thetransmission housing. In this 130 way, the tolerances that must be metby the transmission housing in the area of the housing cover 70 18 andby the clutch device (double-clutch 12) are not as strict. Bearingswhich serve as both axial and radial bearing support are preferablyused. Reference is had to the bearings 68 in the embodiment examples ofFIGS. 1, 3 and 11. The axial and radial bearings, which may be calledcompact bearings depending on the construction, can be constructed so asto allow the cooling fluid, in this case, the cooling oil, to flowthrough them and accordingly enable the advantageous supply of oilbetween the ring part 66 on the one hand and the transmission inputshafts 22, 24 on the other hand.

A further aspect relates to the guiding of the actuating pistons 110 and130. As has already been described in connection with the embodimentexample in FIG. 1, the actuating piston 110 of the first plate clutcharrangement 64 having the radial outer plate stack 76 is guided so as tobe displaceable at the first outer plate carrier 62 and at the secondouter plate carrier 70. This twofold guiding at both the first andsecond outer plate carrier is especially useful particularly when theactuating piston, as in the embodiment examples shown here, acts at theplate stack 76 by a portion 230 (FIG. 2) which projects radially outwardrelatively far from the radial area of the first pressure chamber 118and which accordingly has a relatively long effective lever arm. Thecounterforces of the plate stack exerted on the actuating piston 110 viathe lever arm 230 can accordingly be safely carried off into the outerplate carrier without deformation of the actuating piston 110, whichcould lead to self-locking. As regards the second actuating piston 130,deformations of this kind are of less concern when—as in the embodimentexamples shown herein—the portion of the actuating piston 130 projectingtoward the second plate stack 74 projects less far radially andconsequently no significant “force amplification” occurs due to aneffective lever arm. An additional guiding of the second actuatingpiston 130 corresponding to the guiding of the first actuating piston110 at the second outer plate carrier 70 is achieved likewise with theintermediary of the seal 136 at the wall portion 132 (compare FIG. 1).

An important aspect is the sealing of the pressure chambers and thepressure compensation chambers. With respect to pressure compensationchamber 142, an extremely advantageous construction of the sealingelement 136 is realized in the embodiment example in FIG. 2. The sealingelement 136 is constructed as a curved sealing element 136′ which isplaced over the plate part forming the wall 132 at the radial outer edgeor is injection-molded on this edge. This construction of the sealingelement 136′ facilitates assembly in particular and, as a result, thesealing element 136′ is secured axially to the edge of the wall portion132, that is, it does not move along with the actuating piston 130.

The sealing element 136′ in FIG. 2 can have an axial dimensioning suchthat it acts at an associated portion of the second actuating piston 130when the second plate clutch arrangement 72 is engaged and acts as aspring element which reinforces the opening of the second plate clutcharrangement 72, that is, which pretensions the actuating piston 130 inthe direction of a release position. Also, the seal 114 acting betweenthe second outer plate carrier 70 and the first actuating piston 110 canbe constructed in a corresponding manner, so that the releasing movementof the first actuating piston 110 is also reinforced by the seal 114.Concerning the second actuating piston 130, its releasing movement canalso be supported alternatively or additionally by the wall portion 132which can be constructed so as to be elastically deformable for thispurpose. By reinforcing the releasing movements of the actuating piston,the plate clutch arrangements respond quicker as regards disengagementthan would be the case if only the diaphragm springs 146 and 148(FIG. 1) were provided. With reference to FIG. 2, both diaphragm springsare arranged in the respective pressure compensation chamber 120 or 142.

As an alternative to the construction of the sealing elements as ringelements extending essentially in axial direction in cross section,FIGS. 7c and 7 d show the alternative constructions of the double-clutch12 in the area marked ‘x’ in FIG. 7a. According to the constructionalvariants shown in FIG. 7c, annular grooves 240 are worked into the outerplate carrier 62 (and/or alternatively or additionally—into the piston110); these annular grooves 240, together with an associated surface ofthe other respective part (piston or outer plate carrier), form alabyrinth seal. Sealing elements made of plastic, rubber or the like canthen be dispensed with. This is particularly advantageous inasmuch asthe two elements engaging with one another in a sealing manner can havethe same thermal expansion coefficient. In this way, no substantialchanges in the friction between the elements in sealing engagement withone another and no noteworthy deterioration in sealing action, possiblyleakage, occurs in the event of changes or fluctuation in temperature.

Another possibility for constructing the seals is shown in FIG. 7d.Instead of the sealing ring 112 in FIG. 7a, which extends primarily inaxial direction in cross section, a sealing ring 112′ is providedaccording to FIG. 7d which extends predominantly in radial direction incross section and which is inserted in a shaped portion 250 of the firstactuating piston 110. The sealing element 112′ acts at an innercircumferential surface of the first outer plate carrier 62 in themanner of a wiper or stripper. The sealing element 112′ is tensionedbetween the inner circumferential surface of the outer plate carrier 62and a base of the shaped portion 250 of the actuating piston 110 in sucha way that the curvature of the sealing element 112′ shown in FIG. 7dresults in the disengaged state of the first plate clutch arrangement64. When the first plate clutch arrangement is engaged, the sealingelement 112′ is relaxed and stretched (in cross section). Accordingly,maximum advantage is taken of the sealing engagement of the sealingelement 112′ in the state shown in FIG. 7d, that is, when the actuatingpiston 110 is in its end position corresponding to a disengaged plateclutch arrangement. On the other hand, in contrast to the constructionshown in FIG. 7d, it is preferable that the sealing engagement of therespective sealing element is made use of to the maximum when the clutchis engaged. For this purpose, instead of sealing element 112′, a sealingelement 112″ shown in a detail in FIG. 7d can be inserted into theshaped portion 250, this sealing element 112″ being curved in theopposite direction to that of sealing element 112′ in the relieved statein which it is not yet inserted. In this way, the sealing element 112″is subjected to increasing stretching and accordingly an increasingsealing engagement by the pressure in the pressure space 118 and by theaxial movement of the actuating piston 110 for the purpose of engaging.A stretched tension state of the sealing element 112″, shown as anotherdetail in FIG. 7d, is reached in the course of the engaging movement ofthe first actuating piston 110, possibly not until its axial endengagement position, and can be attributed above all to the influence ofpressure in the pressure chamber 118 on the sealing element 112″ whichadditionally presses the sealing element into the shaped portion 250. Aparticularly effective sealing of the pressure chamber 118 isaccordingly achieved, specifically, in the engaged state above all or inthe course of the engagement of the associated plate clutch arrangement64. It is extremely useful to provide maximum sealing action in thestate of the actuating piston in which it occupies its axial endengagement position, that is, when the plate stack 76 is compressed to amaximum and maximum pressure prevails in the pressure chamber 118. Ifpossible, leakage should not occur especially in this operatingsituation.

A further advantage of the possible construction shown in FIG. 7d forarea x in FIG. 7a (the same applies for the rest of the seals associatedwith the actuating piston) is chiefly that axial installation space iseconomized on because a one-sided groove is sufficient and the groovedepth can extend in a radially extending portion of the actuating piston110 (or, alternatively, of the outer plate carrier). Accordingly, smallwall thicknesses are possible. The groove forming the shaped portion canbe produced simply, for example, by rolling.

The type of arrangement of the actuating piston and especially of theseals associated with it has an effect on the axial and radialinstallation space needed. An important parameter in this connection isthat of the angles α₁, α₂ and α₃ shown in FIG. 5 which amount toapproximately 55° (α₁), approximately 45° (α₂) and approximately 25°(α₃) in the embodiment example in FIG. 5. The angles between ahorizontal line parallel to axis A and the straight lines intersectingseals 114 and 136, seals 112 and 134 and seals 116 and 138 are definedas angles α₁, α₂ and α₃. It has been shown that an arrangement of theseals in an angular area corresponding to an angle α₁, α₂ or α₃ ofapproximately 10° to 70° is advantageous with respect to the compactnessof the double-clutch 12. Angles α₁ and α₂ are particularly important inthis respect. It is apparent from FIG. 5 that it is not required thatseals corresponding to one another must run on the same diameter orradius. Rather, it can be extremely advantageous, for instance, withrespect to compactness, to arrange these seals on different diameters orradii (radii r₁ and r₂ associated with seals 116 and 138 are indicatedin FIG. 5). This can also be contributed to in particular in that theeffective piston surface of the first actuating piston 110 is smallerthan the effective piston surface of the second actuating piston 130 sothat the actuating pressures occurring in the pressure chambers 118 and140 are adapted to one another. This is because, as a rule, both clutcharrangements must transmit the same torque, but the second plate clutcharrangement requires a greater contact pressing force for this purposebecause the average friction radius of its plate stack 74 is smallerthan the plate stack 76 of the first plate clutch arrangement 64.Another possibility for providing the second actuating piston 130 with agreater effective pressure surface subjected to the pressure medium inthe pressure chamber than the first actuating piston 110 is shown inFIG. 13. Additionally, reference is had to the remarks pertaining to theembodiment example in FIG. 1.

Regardless of the construction of the clutch device in particular, it isimportant in wet clutch arrangements to prevent unwanted effects of thecooling fluid, especially the utilized cooling oil or the like.Accordingly, as was already mentioned with respect to the embodimentexample in FIG. 1, unwanted effects of the centrifugal force pressure ofthe oil can be reduced by openings (such as bore holes) in the platecarriers and/or actuating pistons. Accordingly, deformations of theplate carriers in particular, which can result in checking or impairmentof the piston movement, can also be prevented. With respect to providingopenings 162 and 164 in the piston 110 and in the outer plate carrier 62(compare FIG. 11), the construction of the adjacent end plate 166 as aconducting element with a conducting portion 168 is particularly usefulin order to provide for a sufficient volume flow through the plate stack76 in spite of the flow-off possibility for the cooling oil through theopenings 162 and 164. A corresponding through-flow opening 160 is alsoprovided in the carrier plate 60 in the embodiment example of FIG. 11.In FIG. 11, the openings 162, 164 and 260 are designated in theirentirety as a centrifugal force pressure reduction device 262 of thefirst plate clutch arrangement 64.

In the embodiment example of FIG. 13, the first outer plate carrier 62and the first actuating piston 110 are constructed in a special mannerwith respect to the cooling oil outlet openings 162 and 164 so as toeconomize on axial space in the area of the outer plate carrier 72 ofthe second (inner) plate clutch arrangement on the one hand and, ifdesired, to prevent rotation of the first actuating piston 110 relativeto the outer plate carrier 62 on the other hand. For this purpose, thefirst outer plate carrier 62 and the first actuating piston 110 arepartially recessed alternately in the circumferential direction, so thatlocations of the actuating piston 110 that are not recessed engage inrecessed locations of the outer plate carrier 62 and locations of theouter plate carrier 62 that are not recessed engage in recessedlocations of the actuating piston 110. It is useful to provide theaforementioned protection against rotation insofar as additional loadingof the seals acting between the outer plate carrier 62 and the actuatingpiston 110 through microrotation due to engine unevenness can beprevented. In order to achieve this protection against rotation, theactuating piston 110 and the outer plate carrier 62 must also engagewith one another in the engaged state of the first plate clutcharrangement 64, which would not be necessary otherwise.

Concerning the centrifugal force pressure compensation achieved at theactuating piston itself by the pressure compensation chambers, thepressure chamber associated with an actuating piston and the pressurecompensation chamber associated with this actuating piston extend alongthe same radial area in the embodiment examples in FIGS. 2 to 14, sothat fill level limiting means, for example, in the form of a fill levellimiting opening 156 of the pressure compensation chamber 142 of theembodiment example in FIG. 1, are not required. In general, it should benoted with respect to the centrifugal force compensation at the pistonthat the pressure chamber seals and the pressure compensation chamberseals need not necessarily have the same radius. It matters only thatthe pressure difference between the pressure chambers and the associatedcentrifugal force pressure compensation chambers caused by centrifugalforce does not exceed a maximum value and preferably approaches zero.Apart from the outer diameter of the piston chambers which is determinedby the radial outer seals, the pressure difference also depends on theinner diameter of the piston chambers determined by the radial innerseals and can accordingly be influenced by means of this. The fillinglevel limiting means already mentioned can be provided in addition ifrequired.

An important subject pertains to controlling the output losses occurringin the multiple-clutch device, or double-clutch device, as the case maybe, in frictional engagement operating situations of a respective clutcharrangement, especially also in the case of slip operation of the clutcharrangement. For this purpose, it is extremely useful to form the clutcharrangements as wet-type plate clutch arrangements as is the case in theembodiment examples of FIGS. 1 to 14. For an effective through-flowthrough the plate stacks 74 and 76 and, therefore, for an effectivedissipation of friction heat, through-openings which are associated withthe respective plate stack and designated in their entirety by 270 inFIGS. 3 and 4 are preferably provided in the plate carriers. Withrespect to plate stacks which have metal plates (usually steel plates)without facings and plates with facings, the through-openings 270 arepreferably arranged in such a way that the cooling fluid, in this case,the cooling oil, flows directly past the steel plates at least in theengaged state of the respective plate clutch arrangement. This appliesespecially when insulating materials such as paper material are used asfriction facings because then virtually the entire heat capacity of theplate stack is provided by the steel plates.

It is not necessary for the through-openings 270 in the respective innerplate carrier 82 or 86 and the through-openings in the outer platecarrier 62 or 70 to be located directly opposite one another or, as thecase may be, to be aligned. Rather, the flow path of the cooling oilbetween the inner plate carrier and the outer plate carrier is advisablylengthened by an axial displacement of the through-openings relative toone another, so that the oil remains in the area of the plate stacklonger and has more time for heat absorption of the steel plates andfrom the shear gap between plates which can be brought into frictionalengagement with one another.

In this connection, it should be noted that it is particularly advisablewhen the oil flowing through the plate stacks has a releasing effect onthe plates and accordingly reinforces a fast release of the respectiveplate clutch arrangement. For this purpose, an effective oil flowbetween the plate stack and the axially extending ring portion of theouter plate carrier 62 and 70, respectively, and/or the inner platecarrier 82 and 86, respectively, which exerts a dragging effect on theplates is preferably made use of by means of corresponding arrangementsof the through-openings 270 and by providing the possibility for the oilto flow axially out of the area of the plate stack in the direction ofthe actuating piston (in connection with impeding or suppressing anaxial flow of oil out of the area of the plate stack in the oppositedirection toward the carrier plate 60).

The majority of output losses occur during startup at the clutcharrangement which is used as starting clutch. Therefore, it must beensured that the clutch arrangement serving as starting clutch is cooledin a particularly effective manner. If the first plate clutcharrangement 64 having the radial outer plate stack 76 serves as startingclutch, as is preferred, then it is advisable to guide a large portionof the oil volume flow past the inner clutch arrangement 72. For thispurpose, as is shown in FIGS. 4 to 11, the second inner plate carrier 86can be formed with through-openings 280 to enable oil to flow past theplate stack 74 radially outward to the plate stack 76. The inner platecarrier 82 of the outer plate clutch arrangement 64 then preferablyserves as a baffle for the oil flow, so that at least the predominantportion of the oil flowing through the through-openings 280 reaches thethrough-openings 270 in the inner plate carrier 82 which are associatedwith plate stack 76. In this connection, the construction of the endplate 166 with the conducting portion 168 is also particularly useful,since this ensures that the oil flowing to the through-openings 270 inthe inner plate carrier 280 at least predominantly passes through thesethrough-openings and flows through the plate stack 76.

For better control of the friction heat occurring during starting orduring slip operation, the heat capacity of the respective clutcharrangement, especially the first clutch arrangement 64, can beincreased by various measures. Accordingly, it is possible to increasethe number of plates for this clutch arrangement, in this case, thefirst radial outer clutch arrangement, relative to the number of platesin the other clutch arrangement. Accordingly, in the embodiment examplesin FIGS. 2, 11 and 12, the first (outer) clutch arrangement 64 has moreplates than the inner (second) clutch arrangement 72. It was recognizedthat the advantages with respective to the increased heat capacity ofthe plate stack 76 justified the greater input of material, implied bythe different number of plates, for the production of the plates of thetwo clutch arrangements. A further possibility is to produce at leastsome of the friction facings from a heat-conductive material. Forexample, the sintered facings mentioned in connection with theembodiment example in FIG. 1 can be used. Accordingly, for example, inthe embodiment examples in FIGS. 3 to 10 and 13, the axial outer plates(end plates) having the facings, that is, the outer plates on the axialouter side, are outfitted with friction facings of sintered material.Because of the high thermal conductivity of the sintered facings, theseend plates can be effectively utilized for storing output losses,especially starting output losses. These end plates are constructed soas to be comparatively thick axially for a particularly high heatcapacity. Reference is had to the constructions for the embodimentexample in FIG. 1.

A further possibility for increasing the available heat capacityconsists in the use of the carrier plate 60 as a friction surface of theplate stack as is the case in the embodiment examples in FIGS. 2, 11 and12. The carrier plate 60 has a substantially greater mass than anindividual plate and, consequently, a substantially greater heatcapacity and can accordingly temporarily store a larger amount offriction heat. Further, the carrier plate has a large surface at whichit can interact with cooling oil, so that the buffered heat can beeffectively carried off from the carrier plate 60 by the cooling oil.

A difference between the embodiment example in FIG. 11 and theembodiment example in FIG. 12 consists in that the plate which has afacing and which is located on the furthermost right-hand side in theplate stack 76, for example, a paper plate, is shorter in the radialdirection (radially inward) in the embodiment example in FIG. 12 than inthe embodiment example according to FIG. 11. The reason for this step isthat an uneven area pressure of plates having facings can lead toproblems, for example, facing separations. In the case of the embodimentexample in FIG. 11, there is a risk of an uneven area pressure of thefacing-carrying outer plate directly adjacent to the carrier plate 60because the friction surface of the carrier plate associated with theplate passes into a rounded transitional surface region in which theplate is no longer adequately supported axially. Of course, the radialdimensions of the friction surface of the carrier plate could also beincreased to the point that the adjacent plate is uniformly supported atevery point. As a result of this, however, more radial installationspace would be required. On the other hand, the solution in FIG. 12 ispreferred. In this case, the outer plate which is directly adjacent tothe carrier plate 60 and which can be brought into frictional engagementwith the friction surface of the carrier plate 60 is constructed so asto be shorter radially and accordingly has a smaller inner radius thanother outer plates and consequently has a smaller average frictionradius than other outer plates. The radial dimensioning of this outerplate is adapted to the radial dimensioning of the friction surface ofthe carrier plate 60 in such a way that the friction surface of thecarrier plate 60 is substantially flat in the radial area of the outerplate. The rest of the plates (outer plates) having facings can havelarger radial dimensions than the plate with facing (outer plate)directly adjacent to the carrier plate 60 because the adjacent, axiallyoutermost inner plate (steel plate) provides for a uniform area pressureover the larger friction facing surface as well. Other plates withfacings in the plate stack can also differ with respect to their averagefriction radius for making the area pressure more uniform, that is, theycan have somewhat different inner radii in the case of outer plates. Inthis way, temperature profiles which specifically protect against adeformation of the steel plates due to heat can be adjusted in the steelplates not having facings. Further, it is possible by means ofcorresponding temperature profiles to deliberately adjust deformationsof steel plates caused by heat which compensate for deformations ofother steel plates caused by heat, so that the area pressure is rendereduniform overall.

As concerns providing friction facings of different material in a platestack, it has already been noted in connection with the embodimentexample in FIG. 1 that the curve of the coefficient of friction can beadjusted in this way between progressive, neutral and degressive. Aprogressive or at least neutral curve of the coefficient of friction ispreferred in order to counter a buildup of torsional vibrations in thedrivetrain and, to this extent, torsional vibrations do not pose aproblem because, for example, special steps have been taken to damp orsuppress torsional vibrations. Accordingly, it is certainly possible toproduce all of the friction facings of a plate stack from sinteredmaterial so that all of the plates having friction facings, with theirheat capacity, are available as heat buffers.

It has already been mentioned that the two diaphragm springs 146 and 148(compare FIG. 2) are arranged in the respective pressure compensationchambers (120 and 142) in the embodiment examples in FIGS. 2 to 12 so asto make good use of the available installation space. According to theembodiment example in FIG. 12, the outer plate carrier 70 has a stepwith height b at the radial outer side of the diaphragm spring 146 whichserves as an end stop for the actuating piston 110. The step height b isadapted to the thickness of the diaphragm spring 146, so that thediaphragm spring is prevented from bending in the direction opposite tothat shown in FIG. 12 due to the actuating piston 110 traveling to theright. Therefore, a flat contact face for the diaphragm spring 46 at theinner plate carrier 70 is not required, so that the inner plate carrier70 can be designed with respect to its cross-sectional shape in a usefulmanner for purposes of minimizing the required installation space.

In all of the embodiment example of FIGS. 1 to 14, the clutch device iscoupled to the drive unit of the drivetrain via the clutch hub 34, thatis, preferably via a torsional vibration damper as is shown in theexample in FIG. 13. Further, a pump drive shaft 26 is provided as theradially innermost shaft in all of the embodiment examples in FIGS. 1 to14, this radially innermost shaft being coupled to the clutch hub 34 viateeth. In this connection, reference is had to the comments regardingthe embodiment example in FIG. 1.

For technical reasons pertaining to manufacture, the hub is preferablyconstructed in two parts (ring portions 36 and 38 of the hub in FIG. 1).In the embodiment examples in FIGS. 2, 5, 8, 9, 10, 11, 12, 13 and 14,the hub 34 is also constructed in two parts in a corresponding manner,while the hub 34 is constructed in one piece in the embodiment examplesin FIGS. 3, 4, 6 and 7.

Further for technical reasons pertaining to manufacture, it is preferredthat the hub is constructed as a ring part opening toward the driveunit, so that the internal toothing of the hub associated with the pumpdrive shaft 26 can be cleared easily. The opening of the hub can beadvantageously closed by means of a sealing element, for example, asealing journal 180 corresponding to FIG. 5. The sealing journal 180 canbe centered by the internal toothing of the hub 34 and welded to thehub. Another possibility is realized in the embodiment example in FIG.8. In this case, a closure plate part 290 welded to the hub 34, or, moreprecisely, to the ring portion 36 of the hub, is provided in place of asealing journal or the like and has the external toothing 42 associatedwith the torsional vibration damper (not shown) at a flange portion. Theclosure plate part 290 can have a journal-like portion forself-centering of the plate part 290 at the hub 36. Alternatively oradditionally, the plate part 290 can have a journal-like portion whichserves for mutual centering of the engine shaft and transmission inputshafts. The clutch hub 34 itself can also perform this function. In theembodiment example in FIG. 5, the hub 34 is constructed without anopening in the area of the internal toothing.

It should be added that the possibility of reinforcing a disengagementof the respective plate clutch arrangement mentioned in connection withthe sealing element 136′ and in connection with the flow of cooling oilthrough the plates is advantageous in many respects, for example, whenthe plate clutch arrangement in question is to be operated withregulated slip. Other components of the clutch device which are presentin any case can also act in this way, for example, the wall portion 132which defines the second pressure compensation chamber 142 and which canserve as a spring element pretensioning the associated actuating pistonin the releasing direction, as was already indicated in the preceding.

Further details of the double-clutches 12 according to the differentembodiment examples and especially differences between the variousdouble-clutches will be readily discerned from the Figures by the personskilled in the art.

Thus, while there have been shown and described and pointed outfundamental novel features of the present invention as applied to apreferred embodiment thereof, it will be understood that variousomissions and substitutions and changes in the form and details of thedevices illustrated, and in their operation, may be made by thoseskilled in the art without departing from the spirit of the presentinvention. For example, it is expressly intended that all combinationsof those elements and/or method steps which perform substantially thesame function in substantially the same way to achieve the same resultsare within the scope of the invention. Substitutions of elements fromone described embodiment to another are also fully intended andcontemplated. It is also to be understood that the drawings are notnecessarily drawn to scale but that they are merely conceptual innature. It is the intention, therefore, to be limited only as indicatedby the scope of the claims appended hereto.

We claim:
 1. Multiple-clutch device, for installation in a drivetrain ofa motor vehicle between a drive unit and a transmission, comprising afirst clutch arrangement associated with a first transmission inputshaft of the transmission, and a second clutch arrangement associatedwith a second transmission input shaft of the transmission fortransmitting torque between the drive unit and the transmission, atleast one of the clutch arrangements having an actuating piston defininga pressure chamber for pressure medium actuation of the said one clutcharrangement, the actuating piston dividing the pressure chamber from anassociated centrifugal force pressure compensation chamber holding apressure compensation medium, a restoring spring arrangement forrestoring the actuating piston from an actuating position thereof beingcarried in the centrifugal force pressure compensation chamber, anelastic sealing arrangement is associated with the pressure compensationchamber and a boundary wall of the pressure compensation chamber, saidelastic sealing arrangement being engagable with the actuating piston inactuated position thereof and applying pretension to said actuatingpiston for reinforcing a release movement of said actuating pistonduring disengagement of the clutch arrangement.
 2. Clutch deviceaccording to claim 1, in which the centrifugal force pressurecompensation chamber of the at least one clutch arrangement is connectedto a pressure compensation medium supply for supplying pressurecompensation medium to the centrifugal force pressure compensationchamber during operation of the clutch device.
 3. Clutch deviceaccording to claim 2, in which the pressure compensation medium supplycomprises a separate operating fluid supply, said operating fluid supplysupplying at least one other functional unit of the clutch device. 4.Clutch device according to claim 1, in which the sealing arrangement isa sealing element carried at a boundary wall portion of the pressurecompensation chamber and positioned to engage tightly with said wallportion and said actuating piston.
 5. Clutch device according to claim4, in which the sealing element projects from the boundary wall portionin the direction of a stop face of the actuating piston and, at least inthe engaged state of the clutch arrangement, transmits elastic restoringforce to the actuating piston in a direction of a stroke positioncorresponding to a released state of the clutch arrangement.
 6. Clutchdevice according to claim 5, in which the sealing element isinjection-molded on at one of an edge area of the boundary wall portionand fitted over said edge and an area proximal thereto. 7.Multiple-clutch device, for installation in a drivetrain of a motorvehicle between a drive unit and a transmission, comprising a firstclutch arrangement associated with a first transmission input shaft ofthe transmission, and a second clutch arrangement associated with asecond transmission input shaft of the transmission for transmittingtorque between the drive unit and the transmission, at least one of theclutch arrangements having an actuating piston defining a pressurechamber for pressure medium actuation of the at least one clutcharrangement, the at least one clutch arrangement being constructed as aplate clutch arrangement wherein a radial outer clutch annularlyencloses a radial inner clutch arrangement, the actuating pistonassociated with the radial outer clutch arrangement being guided at anouter plate carrier of the radial outer clutch arrangement, and at anouter plate carrier, of the radial inner clutch arrangement so as to beaxially displaceable to seal the pressure chamber, the actuating pistondefining a centrifugal force pressure compensation chamber receivingpressure compensation medium, wherein the actuating piston is guidedfurther at the plate carriers so as to seal the centrifugal forcepressure compensation chamber, the centrifugal force pressurecompensation chamber associated with the radial outer clutch arrangementbeing defined by the outer carrier of the radial inner clutcharrangement, another actuating piston associated with the radial innerclutch arrangement defines another pressure chamber and is guided to beaxially displaceable at the outer plate carrier of the radial innerclutch arrangement and, at a wall of another centrifugal force pressurecompensation chamber associated with said other actuating piston, and soas to seal the associated pressure chamber and the centrifugal forcepressure compensation chamber which receives a pressure compensationmedium, sealing elements cooperate with the actuating pistons of the atleast one clutch arrangement and said another piston to seal therespective ones of the pressure chambers and the centrifugal forcepressure compensation chambers, said sealing elements being arrangedradially staggered and partially overlapping axially.
 8. Clutch deviceaccording claim 7, in which at least one centrifugal force pressurecompensation chamber extends over a different radial area than itsassociated pressure chamber in such a way that an effective pressureapplication surface of an associated piston on a pressure chamber sidethereof is smaller than an effective pressure application surface of thepiston on a pressure compensation chamber side, a pressure compensationchamber limiting surface of the said associated piston extends fartherradially outward than a pressure chamber limiting surface of the saidassociated piston, there being fill level limiting means in the pressurecompensation chamber to limit the filling of the pressure compensationchamber with pressure compensation medium to a maximum radial fillinglevel thereby to prevent centrifugal force overcompensation therein. 9.Clutch device according to claim 8, in which the fill level limitingmeans comprise at least one pressure compensation medium through-openingin a radially extending wall of the pressure compensation chamber. 10.Multiple-clutch device, for installation in a drivetrain of a motorvehicle between a drive unit and a transmission, which clutch devicecomprises a first clutch arrangement associated with a firsttransmission input shaft of the transmission, and a second clutcharrangement associated with a second transmission input shaft of thetransmission for transmitting torque between the drive unit and thetransmission, in which at least one of the clutch arrangements has anactuating piston defining a pressure chamber for pressure mediumactuation of the said one clutch arrangement, the actuating piston beingguided at one of a wall portion defining the pressure chamber and a wallportion defining an associated pressure compensation chamber so as toseal the pressure compensation chamber, wherein at least one sealingring acts between the respective wall portion and the actuating piston,which sealing ring is secured to one of said actuating piston and saidwall portion and acts on a other of said actuating piston and wallportion with axial play during an engaging movement and during adisengaging movement of the actuating piston such that sealingengagement of the sealing ring with the wall portion and actuatingpiston increases.
 11. Clutch device according to claim 10, in which thesealing ring engages on one side in an annular groove in one of theactuating piston and the wall portion.
 12. Clutch device according toclaim 11, in which the sealing ring has at least one portion which actsat one of a surface region of the actuating piston and of the wallportion which extends in an engaging-disengaging direction of theactuating piston and which is planar in this direction.
 13. Clutchdevice according to claim 12, in which the at least one portion of thesealing ring acts at the surface region in the manner of a stripper. 14.Clutch device according to claim 12, in which the sealing ring is curvedin an axial direction in a state of lower tension with a looser sealingengagement with respect to a section plane co-directional with an axisof rotation (A) of the clutch device and is moved to a state of highertension with a tighter sealing engagement during engaging and duringdisengaging movements of the actuating piston, wherein said sealing ringis stretched with respect to said section plane or less sharply curvedin an axial direction.
 15. Clutch device according to claim 14, in whichthe sealing ring is acted upon to the maximum sealing engagement in theengaged state of the respective clutch arrangement.
 16. Clutch deviceaccording to claim 14, in which the wall portion defining the pressurechamber is an outer plate carrier, of the clutch arrangement, saidclutch arrangement being a plate clutch arrangement.
 17. Multiple-clutchdevice, for installation in a drivetrain of a motor vehicle between adrive unit and a transmission, the clutch device comprising a firstclutch arrangement associated with a first transmission input shaft ofthe transmission, and a second clutch arrangement associated with asecond transmission input shaft of the transmission for transmittingtorque between the drive unit and the transmission, each of the clutcharrangements having an actuating piston defining an associated pressurechamber for pressure medium actuation of the clutch arrangement inwhich: (a) a first radial outer sealing element seals the first clutcharrangement actuating piston associated pressure chamber on at least oneof a radial outer side thereof and axially, and which acts between thesaid first clutch arrangement actuating piston and a wall of thepressure chamber associated therewith, and a second radial outer sealingelement seals the second clutch arrangement actuating piston associatedpressure chamber on at least one of a radial outer thereof side andaxially, and which acts between said second clutch arrangement actuatingpiston and a wall of the pressure chamber associated therewith, arearranged at different radial distances from an axis of rotation (A) ofthe clutch device, and (b) a third radial inner sealing element sealsthe first clutch arrangement actuating piston associated pressurechamber on at least one of a radial inner side thereof and axially, andwhich acts between the said first clutch arrangement actuating pistonand a wall of the pressure chamber associated therewith, and a radialfourth inner sealing element which seals the second clutch arrangementactuating piston associated pressure chamber on at least one of a radialinner side thereof and axially, and which acts between said secondclutch arrangement actuating piston and a wall of the pressure chamberassociated therewith, are arranged at different radial distances fromsaid axis of rotation (A) of the clutch device, the first and secondouter sealing elements and the third and fourth inner sealing elementsare so arranged with respect to a section plane co-directional with theaxis of rotation (A) of the clutch device, that one angle α between astraight line which intersects, said first and third sealing elementsand a line parallel to the axis of rotation (A) of the clutch device,and another angle α between a straight line which intersects said secondand fourth sealing elements and a line parallel to the axis of rotation(A) of the clutch device are formed in the disengaged and engaged statesof the two clutch arrangements, the angles α being between about 10° toabout 70°, the actuating pistons divide the respective associatedpressure chambers from an associated pressure compensation chamber whichreceives a pressure compensation medium, wherein (c) a fifth radialouter sealing element seals the pressure compensation chamber of thefirst clutch arrangement on the radial outer side and axially and whichacts between the actuating piston of said first clutch arrangement, anda wall of the associated pressure compensation chamber, and a sixthradial outer sealing element which seals the pressure compensationchamber of the second clutch arrangement on at least one of the radialouter side and axially and which acts between the actuating piston ofsaid second clutch arrangement and a wall of the associated pressurecompensation chamber, said fifth and sixth sealing elements beingarranged at different radial distances from said axis of rotation (A) ofthe clutch device, and wherein (d) a seventh radial inner sealingelement seals the pressure compensation chamber of the first clutcharrangement on at least one of the radial inner side and axially andwhich acts between the actuating piston of said first clutch arrangementand a wall of the associated pressure compensation chamber and an eighthradial inner sealing element seals the pressure compensation chamber ofthe second clutch arrangement on at least one of the radial inner sideand axially and which acts between the actuating piston and of saidsecond clutch arrangement a wall of the pressure compensation chamber,said seventh and eighth sealing elements being arranged at differentradial distances from said axis of rotation (A) of the clutch device.18. The clutch device according to claim 17, in which the angle α isbetween about 20° to about 50°.
 19. The clutch device according to claim17, in which the angle α is between about 30° to about 40°.
 20. Clutchdevice according to claim 17, in which the fifth and sixth sealingelements and the seventh and eighth sealing elements are so arrangedwith respect to a section plane co-directional with the axis of rotation(A) of the clutch device that respective straight lines which intersect,respectively, said fifth and seventh sealing elements, and said sixthand eighth sealing elements, encloses about, with an axis of rotation(A) of the clutch device, an angle α of between about 10° to about 70°,in the disengaged and engage state of the two clutch arrangements. 21.Clutch device according to claim 20, in which the first and the fifthsealing element are separate sealing elements which are arranged at anessentially identical radial distance from the axis of rotation (A) ofthe clutch device, the second and the sixth sealing elements beingseparate sealing elements which are arranged at an essentially identicalradial distance from the axis of rotation (A) of the clutch device, thethird and the seventh sealing elements are formed by a sealing elementassociated with both the pressure chamber and the pressure compensationchamber of the first clutch arrangement, and the fourth and the eighthsealing elements are formed by a sealing element associated with boththe pressure chamber and the pressure compensation chamber of the secondclutch arrangement.
 22. The clutch device according to claim 20, inwhich the angle α is between about 30° to about 60°.
 23. The clutchdevice according to claim 20, in which the angle α is between about 40°to 50°.
 24. Multiple-clutch device, for installation in a drivetrain ofa motor vehicle between a drive unit and a transmission, which clutchdevice comprises a first clutch arrangement associated with a firsttransmission input shaft of the transmission, and a second clutcharrangement associated with a second transmission input shaft of thetransmission for transmitting torque between the drive unit and thetransmission, in which at least one of the clutch arrangements has anactuating piston defining a pressure chamber for pressure mediumactuation of the said one clutch arrangement, the actuating piston beingguided at one of a wall portion defining the pressure chamber and a wallportion defining an associated pressure compensation chamber so as toseal the pressure compensation chamber, wherein a labyrinth sealcomprising at least one annular groove in a surface portion of one ofthe actuating piston and the wall portion, which surface portion extendsin an engaging-disengaging direction of the actuating piston, actsbetween the respective wall portion and the piston.
 25. Multiple-clutchdevice, for installation in a drivetrain of a motor vehicle between adrive unit and a transmission, the clutch device comprising a fistclutch arrangement associated with a first transmission input shaft ofthe transmission, and a second clutch arrangement associated with asecond transmission input shaft of the transmission for transmittingtorque between the drive unit and the transmission, each of the clutcharrangements having an actuating piston defining an associated pressurechamber for pressure medium actuation of the clutch arrangement inwhich: (a) a first radial outer sealing element seals the first clutcharrangement actuating piston associated pressure chamber on at least oneof a radial outer side thereof and axially, and which acts between thesaid first clutch arrangement actuating piston and a wall of thepressure chamber associated therewith, and a second radial outer sealingelement seals the second clutch arrangement actuating piston associatedpressure chamber on at least one of a radial outer thereof side andaxially, and which acts between said second clutch arrangement actuatingpiston and a wall of the pressure chamber associated therewith, arearranged at different radial distances from an axis of rotation (A) ofthe clutch device, and (b) a third radial inner sealing element sealsthe first clutch arrangement actuating piston associated pressurechamber on at least one of a radial inner side thereof and axially, andwhich acts between the said first clutch arrangement actuating pistonand a wall of the pressure chamber associated therewith, and a radialfourth inner sealing element which seals the second clutch arrangementactuating piston associated pressure chamber on at least one of a radialinner side thereof and axially, and which acts between said secondclutch arrangement actuating piston and a wall of the pressure chamberassociated therewith, are arranged at different radial distances fromsaid axis of rotation (A) of the clutch device the first and secondouter sealing elements and the third and fourth inner sealing elementsare so arranged with respect to a section plane co-directional with theaxis of rotation (A) of the clutch device, that one angle α between astraight line which intersects, said first and third sealing elementsand a line parallel to the axis of rotation (A) of the clutch device,and another angle α between a straight line which intersects said secondand fourth sealing elements and a line parallel to the axis of rotation(A) of the clutch device are formed in the disengaged and engaged statesof the two clutch arrangements, the angles a being between about 10° toabout 70°, the actuating pistons divide the respective associatedpressure chambers from an associated pressure compensation chamber whichreceives a pressure compensation medium, wherein (c) a fifth radialouter sealing element seals the pressure compensation chamber of thefirst clutch arrangement on the radial outer side and axially and whichacts between the actuating piston of said first clutch arrangement, anda wall of the associated pressure compensation chamber, and a sixthradial outer sealing element which seals the pressure compensationchamber of the second clutch arrangement on at least one of the radialouter side and axially and which acts between the actuating piston ofsaid second clutch arrangement and a wall of the associated pressurecompensation chamber, said fifth and sixth sealing elements beingarranged at different radial distances from said axis of rotation (A) ofthe clutch device, and wherein (d) a seventh radial inner sealingelement seals the pressure compensation chamber of the first clutcharrangement on at least one of the radial inner side and axially andwhich acts between the actuating piston of said first clutch arrangementand a wall of the associated pressure compensation chamber and an eighthradial inner sealing element seals the pressure compensation chamber ofthe second clutch arrangement on at least one of the radial inner sideand axially and which acts between the actuating piston and of saidsecond clutch arrangement a wall of the pressure compensation chamber,said seventh and eighth sealing elements being arranged at differentradial distances from said axis of rotation (A) of the clutch device,the first and the fifth sealing element are separate sealing elementswhich are arranged at an essentially identical radial distance from theaxis of rotation (A) of the clutch device, the second and the sixthsealing elements being separate sealing elements which are arranged atan essentially identical radial distance from the axis of rotation (A)of the clutch device, the third and the seventh sealing elements areformed by a sealing element associated with both the pressure chamberand the pressure compensation chamber of the first clutch arrangement,and the fourth and the eighth sealing elements are formed by a sealingelement associated with both the pressure chamber and the pressurecompensation chamber of the second clutch arrangement.